Bearing Housing and Assembly of a Screw Compressor

ABSTRACT

An improved bearing housing of a rotary screw compressor is described. The bearing housing is generally shorter than a convention bearing housing. The bearing housing can be configured to enclose and support radial bearings of the screw compressor. The bearing housing can be configured not to enclose axial bearings of the screw compressor in an axial direction.

FIELD

The disclosure herein relates to a rotary type compressor, such as arotary screw compressor, which can be used in, for example, a heating,ventilation, and air-conditioning (“HVAC”) system. More specifically,the disclosure relates to a bearing housing of a rotary screw compressorto support and enclose discharge axial bearings. The bearing housingherein can improve location of the rotors which may result in improvedcompressor performance and reliability, and machining capability of thehousing may also be improved.

BACKGROUND

A screw compressor is a type of positive displacement compressor thatcan be used to compress various working fluids, such as for examplerefrigerant vapor. Such screw compressors may be used in refrigerationunits, such as for example, water chillers as part of a HVAC system. Thescrew compressor typically includes one or more rotors that rotaterelative to bearings such as for example radial and axial bearings atthe discharge end. A bearing housing and cover are often part of theassembly of the screw compressor to enclose and support the bearings,e.g. radial and axial bearings. During operation, the bearing cover canhave a discharge outlet or port so that a compressed working fluid (e.g.refrigerant vapor) can be discharged from an axial end of the rotors andout of the bearing housing and cover.

SUMMARY

According to an embodiment of the present invention, a variable capacityscrew compressor comprises a rotor housing, a motor, and a variablespeed drive. The rotor housing comprises a suction port, a workingchamber, a discharge port, and at least two screw rotors that comprise afemale screw rotor and a male screw rotor being positioned within theworking chamber for cooperatively compressing a fluid. The suction port,the at least two screw rotors and the discharge port are configured inrelation to a selected rotational speed. The selected rotational speedoperates at least one screw rotor at an optimum peripheral velocity thatis independent of a peripheral velocity of the at least one screw rotorat a synchronous motor rotational speed for a rated screw compressorcapacity. A motor is operable to drive the at least one screw rotor at arotational speed at a full-load capacity that is substantially greaterthan the synchronous motor rotational speed at the rated screwcompressor capacity. A variable speed drive receives a command signalfrom a controller and generates a control signal that drives the motorat the rotational speed.

In another embodiment, a method for sizing at least two screwcompressors is provided. The target capacity for each screw compressoris selected. Each screw compressor has a different rated capacity andfurther comprises a suction port, a working chamber, a discharge port,and at least two screw rotors being positioned within the workingchamber for cooperatively compressing a fluid. The rotational speed isselected to operate at least one screw rotor in each screw compressor atan approximately constant optimum peripheral velocity that isindependent of the rated capacity of each screw compressor. The suctionport, the at least two screw rotors and the discharge port areconfigured together with the rotational speed for each screw compressor.

In another embodiment, a refrigeration chiller, having at least onerefrigeration circuit, comprises a variable capacity screw compressor,condenser, expansion valve and evaporator. The variable capacitycompressor comprises a rotor housing, a motor housing and a variablespeed drive. The rotor housing further comprises a suction port, aworking chamber, a discharge port, and at least two screw rotors thatcomprise a female screw rotor and a male screw rotor being positionedwithin the working chamber for cooperatively compressing a fluid. Thesuction port, the at least two screw rotors and the discharge port areconfigured in relation to a selected rotational speed. The selectedrotational speed provides at least one screw rotor to operate at anoptimum peripheral velocity that is independent of a peripheral velocityof the at least one screw rotor at a synchronous motor rotational speedfor a rated screw compressor capacity. The motor housing furthercomprises a motor, the motor is operable to drive the at least one screwrotor at a rotational speed at a full-load capacity that issubstantially greater than the synchronous motor rotational speed at therated screw compressor capacity. The variable speed drive is configuredto receive a command signal from a controller and to generate a controlsignal that drives the motor at the rotational speed. A condenser iscoupled to the discharge port of the variable capacity screw compressor.The condenser is configured to cool and condense fluid received from thedischarge port. An expansion valve is coupled to the condenser. Theexpansion valve is configured to evaporate at least a portion of fluidreceived from the condenser by lowering pressure of fluid received fromthe condenser. An evaporator is coupled to the expansion valve. Theevaporator is configured to evaporate fluid received from the expansionvalve and to provide fluid to the suction port of the variable capacityscrew compressor.

An improved bearing housing of a rotary screw compressor is described. Abearing housing is generally configured to suitably enclose and supportdischarge radial bearings which are located at a discharge side of thecompressor, for example toward the axial end of the rotors.

In previously known designs, the discharge bearing housing of a screwcompressor is constructed to be a relatively long part, which enclosesand/or supports the discharge radial bearings, the axial bearings, andthe bearing retaining assembly for example the axial bearing retainers.

In the bearing housing shown and described herein, a shorter bearinghousing suitably encloses and/or supports the discharge radial bearings,but does not enclose or support the axial bearings and retainingassembly, e.g. the axial bearing retainers. A bearing cover is providedwhich encloses the axial bearings and the axial bearing retainers. Theshorter bearing housing may be simpler to fabricate and the accuracy ofthe discharge axial bearing bores may be improved due to shorter reachesand shorter machine tool boring bars, compared to a convention bearinghousing. Due to the new bearing housing design, the bearing cover may befabricated relatively easier using machine tools that may not be asprecise as those used to fabricate the bearing housing. That is, thedesign of the bearing housing herein can improve for example themachining capability of the housing, such as for example by enablingshort machine cutter tooling and short reaches for the machining centerof the housing. The shorter bearing housing can improve location of therotors such as during assembly, such as for example by improving theaccuracy of the discharge bearing bores, which may result in improvedcompressor performance and reliability.

In one embodiment, a bearing assembly may include a bearing cover and abearing housing. The bearing housing includes a cavity that isconfigured to enclose and/or support a discharge radial bearing. Thecavity has a depth. The discharge radial bearing has a length. The depthof the cavity may be configured to be no more than the length of thedischarge radial bearing so that the cavity can be configured to encloseand/or support the discharge radial bearing, but not the axial bearing.

Other features and aspects of the embodiments will become apparent byconsideration of the following detailed description and accompanyingdrawings.

BRIEF DESCRIPTION OF THE DRAWINGS

Reference is now made to the drawings in which like reference numbersrepresent corresponding parts throughout.

FIG. 1 illustrates an embodiment that incorporates a screw compressorarranged as part of a refrigeration chiller system.

FIG. 2 illustrates a cross sectional view of a screw compressoraccording to one embodiment.

FIG. 3 illustrates an additional cross sectional view of a screwcompressor according to one embodiment.

FIG. 4 illustrates an embodiment of a refrigeration chiller andcontroller system according to one embodiment.

FIG. 5 illustrates a partial sectional view of a screw compressor, withwhich the embodiments as disclosed herein can be practiced.

FIG. 6 illustrates an enlarged sectional view of an area A of the screwcompressor in FIG. 5.

FIG. 7 illustrates a conventional bearing housing.

DETAILED DESCRIPTION

As a preface to the detailed description, as used in this specificationand the appended claims, the singular forms “a,” “an,” and “the” alsoinclude plural referents, unless the context clearly dictates otherwise.References in this specification to “one embodiment,” “an embodiment,”“an example embodiment,” etc., indicate that the described embodimentmay include a particular feature, structure, or characteristic; however,every embodiment may not necessarily include the particular feature,structure, or characteristic. When a particular feature, structure, orcharacteristic is described in connection with an embodiment, otherembodiments may incorporate or otherwise implement such feature,structure, or characteristic whether or not explicitly described.

Referring now to FIGS. 1-4, components of a chiller or chiller system 10are illustrated. Chiller 10 includes many other conventional featuresnot depicted for simplicity of the drawings.

Chiller system 10 is directed to refrigeration systems. Chiller 10 is inthe range of about 20 to 500 tons or larger, particularly where therefrigeration system includes a multiple stage compressor arrangement.Persons of ordinary skill in this art will readily understand thatembodiments and features of this invention are contemplated to includeand apply to, not only single stage compressors/chillers, but also to(i) multiple stage compressors/chillers and (ii) single and/ormultistage compressor/chillers operated in parallel.

As shown, chiller 10 comprises a screw compressor system 12 (alsosometimes referred to as a screw compressor 12), a condenser 14, and anevaporator 20, all of which are serially connected to form a semi- orfully-hermetic, closed-loop refrigeration system. Chiller 10 maycirculate a fluid 80 (such as, for example, a refrigerant) to controlthe temperature in a space such as a room, home, or building. The fluid80 may be circulated to absorb and remove heat from the space and maysubsequently reject the heat elsewhere.

Fluid 80 may be a refrigerant. The refrigerant may be selected from anazeotrope, a zeotrope or a mixture or blend thereof in gas, liquid ormultiple phases. For example, such refrigerants may be selected from:R-123, R-134a, R-1234yf, R-410A, R-22 or R-32. Because embodiments ofthe present invention are not restricted to the refrigerant chosen,embodiments of the present invention are also adaptable to a widevariety of refrigerants that are emerging, such as low global warmingpotential (low-GWP) refrigerants.

FIG. 1 illustrates the condenser 14. Condenser 14 is shown as a shelland tube flooded-type. The condenser 14 can be arranged as a singleevaporator or multiple evaporators in series or parallel, e.g.connecting a separate or multiple evaporators to each compressor.Condenser 14 may include condenser tubing 16. Fluid 80 may pass acrossthe condenser tubing 16 through which cool air or cool liquid flows.

Condenser 14 may be fabricated from carbon steel and/or other suitablematerial, including copper alloy heat transfer tubing. Condenser tubing16 can be of various diameters and thicknesses, and comprised typicallyof copper alloy. In addition, condenser tubing 16 may be replaceable,mechanically expanded into tube sheets and externally finned seamlesstubing. Other known types of condenser 14 are contemplated.

Condenser 14 may be configured to communicate fluid 80 from a dischargepassage 36. Discharge passage 36 may be configured to receive the fluid80, or may be coupled to the condenser 14 through an oil separator 24,as depicted in FIG. 1. Other configurations are contemplated. The oilseparator 24, when employed, separates oil from the fluid 80 and returnsthe oil via an oil supply passage 26 to the screw compressor 12 forreuse. The oil may be reused to, for example, cool the fluid 80, coolscrew rotors 42, seal the interfaces between the screw rotors 42themselves, seal the interfaces between the screw rotors 42 and thewalls of a working chamber 44, and/or lubricate bearings 46, 48.

Condenser 14 may transform the fluid 80 from a superheated vapor to asaturated liquid. As a result of the cool air or cool liquid passingacross the condenser tubing 16, fluid 80 may reject or otherwise deliverheat from the chiller 10 to another fluid, like air or liquid, in a heattransfer relation, which in turn carries the heat out of the system.

An expansion valve 18 may employed, as shown in FIG. 1. Expansion valve18 may be configured to receive fluid 80 from condenser 14. Fluid 80received from condenser 14 typically is in a thermodynamic state knownas a saturated liquid. The expansion valve 18 may abruptly reduce thepressure of the fluid 80. The abrupt pressure reduction may causeadiabatic flash evaporation of at least a portion of the fluid 80. Inparticular, the adiabatic flash evaporation may result in a liquid andvapor mixture of the fluid 80 that has a temperature that is colder thanthe temperature of the space to be cooled.

Evaporator 20 is shown in FIG. 1 as a shell and tube flooded-type. Theevaporator 20 can be arranged as a single evaporator or multipleevaporators in series or parallel, e.g. connecting a separate ormultiple evaporators to each compressor. Evaporator 20 may includeevaporator tubing 22. Fluid 80 may pass across the evaporator tubing 22through which cool air or cool liquid flows.

Evaporator 20 may be fabricated from carbon steel and/or other suitablematerial, including copper alloy heat transfer tubing. Evaporator tubing22 can be of various diameters and thicknesses, and comprised typicallyof copper alloy. In addition, evaporator tubing 22 may be replaceable,mechanically expanded into tube sheets and externally finned seamlesstubing. Other known types of evaporator 20 are contemplated.

Evaporator 20 is configured, as illustrated in FIG. 1, to receive fluid80 communicated from the expansion valve 18. Fluid 80 received by theevaporator 20 in the refrigeration loop may be relatively colder than itwas when discharged from the screw compressor 12. The oil returnapparatus 28, when employed, separates oil from the fluid 80 and returnsthe oil via an oil return passage 30 to the screw compressor 12 forreuse. The oil may be reused to, for example, cool the fluid 80, coolscrew rotors 42, seal the interfaces between the screw rotors 42themselves, seal the interfaces between the screw rotors 42 and thewalls of a working chamber 44, and/or lubricate the bearings 46, 48.

The evaporator 20 may absorb and remove heat from the space to becooled, and the condenser 14 may subsequently reject the absorbed heatto air or liquid that carries the heat away from the space to be cooled.In operation, warm air or liquid may be circulated from the space to becooled across the evaporator tubing 22. The warm air or liquid passingacross the evaporator tubing 22 may cause a liquid portion of the coldfluid 80 to evaporate. At the same time, the warm air or liquid passedacross the evaporator tubing 22 may be cooled by the fluid 80. It shouldbe understood that any configuration of the condenser 14 and/orevaporator 20 may be employed that accomplishes the necessary phasechanges of fluid 80.

The chilled or heated water is pumped from the evaporator 20 to an airhandling unit (not shown). Air from the space that is being temperatureconditioned is drawn across coils in the air handling unit thatcontains, in the case of air conditioning, chilled water. The drawn-inair is cooled. The cool air is then forced through the air conditionedspace, which cools the space.

Additionally, though not shown, an economizer 32 may be incorporated toinclude an economizer cycle. Economizer 32 or a subcooling cycle (notshown), or both, may be employed in the refrigeration cycle and returnthe fluid 80 to the screw compressor 12 via suction passage 34 or otherpassage (not shown) depending on the configuration required theapplication conditions.

Referring to FIGS. 2 and 3, screw compressor 12 typically comprises arotor housing 40 and an electric motor housing 50. Screw compressor 12may be formed, all or in part, of gray cast iron, for example. Othermaterials may be used to form the screw compressor 12. Screw compressor12, according to embodiments of the present invention, facilitateshighly efficient operation at full-load and part-load conditions over apreselected screw capacity range.

Motor housing 50 houses a motor 52 in an embodiment of the presentinvention. Electric motor 52 may coupled to a variable frequency drive38. The electric motor 52 drives meshed screw rotors 42. Motor housing50 may be integral to the rotor housing 40.

The rotor housing 40 may have a low pressure end and a high pressure endthat each contain a suction port 76 and discharge port 78, respectively.Suction port 76 and discharge port 78 are in open-flow communicationwith the working chamber 44. The suction port 76 and the discharge port78 may each be an axial, a radial or a mixed (a combination of a radialand an axial) port.

The suction port 76 may receive the fluid 80 at a suction pressure and asuction temperature. The suction port 76 may receive fluid 80 fromsuction passage 34 in thermodynamic states known as a saturated vapor ora superheated vapor. The screw compressor 12 may compress the fluid 80as the screw compressor 12 communicates the fluid 80 from the suctionport 76 to the discharge port 78. Fluid 80 passing through the dischargeport 78 discharges into discharge passage 36.

Compressing the fluid 80 may also result in the fluid 80 beingdischarged at a discharge temperature that is higher than the suctiontemperature. The fluid 80 discharged from the discharge port 78 may bein a thermodynamic state known as a superheated vapor. Accordingly,fluid 80 discharged from the screw compressor 12 may be at a temperatureand a pressure at which the fluid 80 may be readily condensed with acooling air or a cooling liquid.

Suction port 76 and discharge port 78 are configured to minimize flowlosses, when at least one of the rotors 42 is operated at anapproximately constant peripheral velocity. The suction port 76 may belocated where fluid 80 exits the suction area of screw compressor 12 andis drawn into the working chamber 44. The suction port 76 may be sizedto be as large as possible to minimize, at least, the approach velocityof the fluid 80. The location of the suction port 76 in the rotorhousing 40 also may be configured to minimize turbulence of fluid 80prior to entry into the rotors 42.

Discharge port 78 may be sized larger than theoretically necessary toprovide a thermodynamic optimum size and thereby, reduce the velocity atwhich the fluid 80 exits the working chamber 44. The discharge port 78may be generally located where fluid 80 exits the working chamber 44 ofscrew compressor 12. The discharge port 78 location in the rotor housing40 may be configured such that the maximum discharge pressure can beattained in the rotors 42 prior to being delivered into the dischargepassage 36. In addition, screw compressor 12 may incorporate a muffler58 or other apparatus suitable for noise reduction.

Referring again to FIG. 3, rotors 42 are mounted for rotation in aworking chamber 44. The working chamber 44 comprises a volume that isshaped as a pair of parallel, intersecting flat-ended cylinders, and isclosely toleranced to the exterior dimensions and geometry of theintermeshed screw rotors 42. The plurality of meshed screw rotors 42 a,42 b may define one or more compression pockets between the screw rotors42 a, 42 b and the interior chamber walls of the rotor housing 40. Therotor housing 40 has little separation from the rotors 42. Milling,machine grinding or molding can be employed to achieve high accuracy andtight tolerances between rotors 42 flutes and lobes and the rotorhousing 40.

First screw rotor 42 a and second screw rotor 42 b are disposed in acounter-rotating, intermeshed relationship and cooperate to compress afluid. At least one of rotors 42 is cooperatively configured with motor52 to be operable at a rotational speed for a screw compressor capacitywithin a preselected screw compressor capacity range. The selectedrotational speed at full-load capacity is substantially greater than asynchronous motor rotational speed at a rated capacity (also referred toherein as rated screw compressor capacity) for screw compressor 12.

In the embodiment illustrated, rotor 42 a may be called a female screwrotor and comprise a female lobed/fluted body or working portion(typically a helical or spiral extending land and groove). Rotor 42 bmay be called a male screw rotor and comprise a male lobed/fluted bodyor working portion (typically a helical or spiral extending land andgroove).

Rotors 42 include shaft portions, which are, in turn, mounted to thehousing of screw compressor 12 by, for example, one or more bearings 46,48. The exemplary bearings 46, 48 will also be configured with tightclearances in relation to at least rotors 42 and rotor housing 40.

Compression of the fluid 80 in screw compressor 12 produces axial andradial forces. The configurations of embodiments of the presentinvention may also mitigate time varying and non-uniform rotor movementsand forces against chamber walls, bearings, and end surfaces of thescrew compressor 12 caused by the interaction of the screw rotors 42 a,42 b, the axial forces, and the radial forces.

As mentioned, a lubricating fluid, typically oil, may be delivered fromoil supply passage 26 or oil return passage 30 to the screw compressor12. The lubricating fluid provides cushioning films for the walls of theworking chamber 44, rotors 42 a, 42 b, and bearings 46, 48 of the screwcompressor 12, but does little to prevent the transmission of the timevarying and non-uniform axial and radial forces. The screw compressor 12may also utilize an expander (not shown), which may also be integral toscrew compressor 12, to recover energy available from the refrigerationcycle as the high pressure liquid expands through the expander to alower pressure.

The electric motor 52 in one exemplary embodiment may drive at least oneof the rotors 42 in response to command signals 62 received from thecontroller 60. The horsepower of preferred motor 52 can vary in therange of about 125 horsepower to about 2500 horsepower. Torque suppliedby the electric motor 52 may directly rotate at least one of the screwrotors 42. Employing motor 52 and variable speed drive 38, screwcompressor 12 of embodiments of the present invention may have a ratedscrew compressor capacity within the range of about 35-tons to about150-tons or more and have a full-load speed range within about 4,000revolutions per minute to about 15,000 revolutions per minute, when thefluid is an R-134a refrigerant.

While conventional types of motors, like induction motors, can be usedwith and will provide a benefit when employed with embodiments of thepresent invention, a preferred motor 52 comprises a direct drive,variable speed, hermetic, permanent magnet motor. Permanent magnet motor52 can increase system efficiencies over other motor types. The choiceof motor 52 may be affected by cost and performance considerations.

Referring to FIGS. 2 and 3, the permanent magnet motor 52 comprises amotor stator 54 and a motor rotor 56. Stator 54 consists of wire coilsformed around laminated steel poles, which convert variable speed drive38 applied currents into a rotating magnetic field. The stator 54 ismounted in a fixed position in the screw compressor 12 and surrounds themotor rotor 56, enveloping the rotor 56 with the rotating magneticfield. Motor rotor 56 is the rotating component of the motor 52 and mayconsist of a steel structure with permanent magnets, which provides amagnetic field that interacts with the rotating stator magnetic field toproduce rotor torque. In addition, permanent magnet motor 52 may beconfigured to receive variable frequency control signals and to drivethe at least two screw rotors per the received variable frequencycontrol signals.

The motor rotor 56 may have a plurality of magnets and may comprisemagnets buried within the rotor steel structure or be mounted at therotor steel structure surface. Motor rotor 56 surface mount magnets aresecured with a low loss filament, metal retaining sleeve or by othermeans to the rotor steel support. Further manufacturing, performance,and operating advantages and disadvantages can be realized with thenumber and placement of permanent magnets in the motor rotor 56. Forexample, surface mounted magnets can be used to realize greater motorefficiencies due to the absence of magnetic losses in interveningmaterial, ease of manufacture in the creation of precise magneticfields, and effective use of rotor fields to produce responsive rotortorque. Likewise, buried magnets can be used to realize a simplermanufactured assembly and to control the starting and operating rotortorque reactions to load variations.

The performance and size of the permanent magnet motor 52 is due in partto the use of high energy density permanent magnets. Permanent magnetsproduced using high energy density magnetic materials, typically atleast 20 MGOe (Mega Gauss Oersted), produce a strong, more intensemagnetic field than conventional materials. With a motor rotor 56 thathas a stronger magnetic field, greater torques can be produced, and theresulting motor 52 can produce a greater horsepower output per unitvolume than a conventional motor, including induction motors. By way ofcomparison, the torque per unit volume of permanent magnet motor 52 canbe at least about 75 percent higher than the torque per unit volume ofinduction motors used in refrigeration chillers of comparablerefrigeration capacity. The result is a smaller sized motor to meet therequired horsepower for a specific compressor assembly.

The permanent magnet motor 52 of an embodiment of the present inventionis compact, efficient, reliable, and relatively quieter thanconventional motors. As the physical size of the screw compressor 12 isreduced, motor 52 used can be scaled in size to fully realize thebenefits of improved fluid flow paths and compressor element shape andsize. Motor 52 is reduced in volume by approximately 30 percent or more,when compared to conventional existing designs for compressor assembliesthat employ induction motors and have refrigeration capacities in excessof 35-tons. The resulting size reduction of embodiments of the presentinvention provides a greater opportunity for efficiency, reliability,and quiet operation through use of less material and smaller dimensionsthan has been achieved through more conventional practices.

Any bearings employed with motor 52 may be rolling element bearings(REB) or hydrodynamic journal bearings. Such bearings may be oillubricated. Oil-free bearing systems may be employed. A special class ofbearing which is refrigerant lubricated is a foil bearing and anotherbearing type uses REB with ceramic balls. Each bearing type hasadvantages and disadvantages that should be apparent to those of skillin the art. Bearings should be selected to facilitate highly efficientoperation of the screw compressor 12 at reduced speeds for capacitymodulation and to minimize rotor dynamics and vibration associated withreduced speeds. Any bearing type may be employed that is suitable ofsustaining rotational speeds in the range of about 2,000 RPM to about20,000 RPM.

The motor rotor 56 and motor stator 54 end turn losses for the permanentmagnet motor 52 are very low compared to some conventional motors,including induction motors. The motor 52, therefore, may be cooled bymeans of fluid 80 (typically, refrigerant). When fluid 80 is employedfor cooling motor 52, fluid 80 may only need to contact the outsidediameter of the stator 54. Cooling the motor 52 in this way allows forthe elimination of the motor cooling feed ring that is typically used ininduction motor stators. Alternatively, refrigerant may be metered tothe outside surface of the stator 54 and to the end turns of the stator54 to cool the motor 52.

In addition, the torque that is needed from motor 52 comes essentiallyfrom the internal pressure distribution in the rotors 42, which is afunction of rotors 42 geometry and the operating conditions. Thatinternal pressure distribution within the rotors 42 provides the loadagainst which the motor 52 has to work. Employing embodiments of thisinvention without a mechanical unloader results in a theoretical torquethat may be essentially constant over a full range of operatingconditions, and for a given operating condition, a ratio of theoreticalto actual torque on the motor 52 that may be approximately constant,despite decay in the actual torque during operation due to changinglosses and leakage, for example. In contrast, for a given operatingcondition, conventional screw compressors invoking a mechanical unloaderwill have significant torque fluctuations or variations over time.

As illustrated in FIG. 4, a variable speed drive 38 may drive the motor52 and in turn, screw compressor 12. The speed of the motor 52 can becontrolled by varying, for example, the frequency of the electric powerthat is supplied to the motor 52. Use of a permanent magnet motor 52 andvariable speed drive 38 moves some conventional motor losses outside ofthe refrigerant loop. The efficiency of the variable speed drive 38,line input to motor shaft output, preferably can achieve a minimum ofabout 95 percent over the system operating range.

The variable speed drive 38 drives the screw compressor 12 at theoptimum, or near optimum, rotational speed at each capacity over thepreselected screw compressor capacity range for a screw compressor 12 ofa given rated capacity. The variable speed drive 38 may be refrigerantcooled, water cooled or air cooled. As mentioned, similar to cooling ofmotor 52, the variable speed drive 38, or portions thereof, may be byusing a refrigerant circulated within the chiller system 10 or by otherconventional cooling means. How the motor 52 and/or variable speed drive38 are cooled should be understood as dependent on the operational andenvironmental conditions in which the motor 52 and/or variable speeddrive 38 reside in operation.

The variable speed drive 38 typically will comprise an electrical powerconverter comprising a line rectifier and line electrical currentharmonic reducer, power circuits and control circuits (such circuitsfurther comprising all communication and control logic, includingelectronic power switching circuits). Conditions in which the screwcompressor 12 is employed may justify employing more than one variablespeed drive 38 for chiller 10.

The variable speed drive 38 can be configured to receive command signals62 from a controller 60 and to generate a control signal 64. Thevariable speed drive 38 will respond, for example, to signals 62received from a microprocessor (also not shown) associated withcontroller system 60 to increase or decrease the speed of the motor 52by changing the frequency of the current supplied to motor 52.Controller 60 may be configured to receive status signals 82 indicativeof an operating point of the screw compressor, and to generate commandsignals that requests the electric motor system to drive the screwcompressor per a preselected operating parameter. Status signals 82 maydeliver similar or different status information depending, for example,on the intended purpose of the sensor selected. Controller 60 maygenerate command signals 62 per a preselected operating parameter, likea torque profile for screw compressor 12. Control signal 64 can drivethe high energy density motor 52 at a rotational speed substantiallygreater than a synchronous motor rotational speed for the rated screwcompressor capacity and drive the motor 52, and in turn at least onescrew rotor 42, at an optimum peripheral velocity independent of therated screw compressor capacity.

The motor 52 and the variable speed drive 38 have power electronics forlow voltage (less than about 600 volts), 50 Hz and 60 Hz applications.Typically, an AC power source (not shown) will supply multiphase voltageand frequency to the variable speed drive 38. The AC voltage or linevoltage delivered to the variable speed drive 38 will typically havenominal values of 200V, 230V, 380V, 415V, 480V, or 600V at a linefrequency of 50 Hz or 60 Hz depending on the AC power source.

By the use of motor 52 and variable speed drive 38, the speed of motor52 can be varied to match varying system requirements. Speed matchingresults in approximately 30 percent more efficient system operationcompared to a compressor without a variable speed drive 38. By runningcompressor 12 at lower speeds when the load on the chiller is not highor at its maximum, sufficient refrigeration effect can be provided tocool the reduced heat load in a manner which saves energy, makes thechiller 10 more economical from a cost-to-run standpoint, andfacilitates highly efficient chiller 10 operation as compared tochillers which are incapable of such load matching at the rotationalspeeds possible via embodiments of the present invention. For example, arated screw compressor capacity of about 100-tons configured accordingto embodiments of the present invention could be efficiently operableover a preselected screw capacity range of about 75-tons to about125-tons.

Screw compressor 12 can be operated at rotational speeds substantiallyhigher than synchronous motor rotational speeds for a given ratedcapacity of the screw compressor 12. The specific optimum speed for therated screw compressor capacity range is a function of screw compressorcapacity and head pressure. Embodiments of the present inventiondramatically improve the discharge porting of fluid 80 and in turn,allow for screw compressor 12 to be operated at a significantlyincreased rotational speed over the rotational speed that gives the bestperformance for conventionally sized rotors and ports. For example, theselected rotational speed for a rated screw compressor capacity of about100-tons, according to embodiments of the present invention, is about5800 revolutions per minute, when the fluid is an R-134a refrigerant. Incontrast, a conventional screw compressor with a rated capacity of about100-tons has a synchronous motor rotational speed is about 3400revolutions per minute, when the fluid is an R-134a refrigerant.

The allowable range of rotational speed for a particular rated capacityof a screw compressor 12 is selected to achieve an optimum peripheralvelocity of at least one of the screw rotors independent of the ratedcapacity of screw compressor 12 that results in a relatively uniformhigh efficiency across the screw compressor product family (e.g.60-tons, 80-tons, 100-tons and 150-tons.) The optimum peripheralvelocity is a constant product of the rotational speed and the radius ofat least one of the rotors 42, typically, the male rotor 42 b. Theapproximately constant optimum peripheral velocity is, for example, inthe range between about 131 feet per second (about 40 meters per second)to about 164 feet per second (about 50 meters per second). In oneembodiment, the approximately constant optimum peripheral velocity isbetween about 42 meters per second (about 137 feet per second) to about45 meters per second (about 147 feet per second) in high pressureapplications, when R-134a refrigerant is the fluid 80. Persons of skillin the art would understand that, for a low pressure application or fora different primary fluid 80, or both, the optimum peripheral velocitymay be different.

The rotational speed of the motor 52 may be selected in combination withconfiguring rotors 42, suction port 76 and discharge port 78 for eachtarget capacity to achieve an approximately constant optimum peripheralvelocity of at least one of the screw rotors 42 regardless of the ratedcapacity of the screw compressor 12. That is, specific combinations ofscrew rotors 42, inlet port 76, discharge port 78 and the operationalrotational speed are selected such that each specific combinationenables each screw compressor 12 to run at approximately the sameoptimum peripheral velocity for each different rated capacity and, inturn, to produce relatively the same high efficiency between or amongeach different rated capacity of screw compressor 12.

Embodiments of the present invention include a method of sizing of atleast two screw compressors 12 with different rated capacities thatachieve approximately constant efficiency across the screw compressorproduct family (e.g. 60-tons, 80-tons, 100-tons and 150-tons.). Byemploying embodiments of this invention, the isoentropic efficiencyversus capacity (in tons) of screw compressor 12 is significantlyincreased, on the order of 15 percent, over a conventional screwcompressor. In addition, because screw compressor 12 is operated atrelatively higher speed, the screw compressor 12 can slowed down on theorder of 20-30 percent of the speed for the operating capacity and stillhave an approximately constant peak efficiency or efficiency plateau ascompared to the efficiency at the rated screw compressor capacity.

The target capacity for each screw compressor 12, each having adifferent rated capacity, is selected. The rotational speed is alsoselected based on the target capacity of each screw compressor 12 tooperate at least one screw rotor 42 in each screw compressor 12 at anapproximately constant optimum peripheral velocity that is independentof the rated capacity of each screw compressor 12. The suction port 76,the at least two screw rotors 42 and the discharge port 78 areconfigured together with the rotational speed selected for each screwcompressor 12.

Specifically, driving screw compressor 12 at an optimum peripheralvelocity allows for each rotor 42 to have a geometry and a profile thatmay remain the same for a wide range of preselected screw compressorcapacities for the rated screw compressor capacity. Each of the rotors42, though, may have a different geometry and a profile for eachdifferent rated screw compressor capacity that will enable at least onescrew rotor to be operated at a selected rotational speed that producesan approximately constant optimum peripheral velocity between or amongeach rated capacity of each screw compressor 12. The volumetric ratio ofthe screw compressor 12 is selected as a function of the loadingconditions in which the screw compressor 12 will be used. By way ofexample, in embodiments of the present invention, more than twovolumetric ratios, potentially four, five or more, are contemplated overa range of rated screw compressor capacities. The volumetric ratio mayalso be such that the system compression ratio and the internalcompression ratio closely match. The rotor 42 profile may be a balanceof the length of the sealing line, flow cross sectional area andblow-hole area size.

The geometry and profile are generally defined, in part, by the numberof lobes in each rotor, the wrap angle, the length of the rotors and thediameter of the rotors, for example. Screw rotor 42 has a profile takenin a plane transverse to the parallel axes of the male rotor 42 b andthe female rotor 42 a. The profile of rotors 42 can be symmetric orasymmetric, and circular, elliptical, parabolic, hyperbolic, forexample. Rack generation of rotors 42 profile may be employed. Selectinga profile of rotors 42 is a balance of the internal leakage path offluid 80 during operation of screw compressor 12 and the portingconfiguration of suction port 76 and discharge port 78, such that screwcompressor 12 has an approximately constant optimum peripheral velocity.

More specifically, for example, at an about 44 m/s optimum peripheralvelocity for at least one rotor 42 of a 100-ton screw compressor, theresulting male rotor 42 b has a wrap angle of about 347 degrees and thefemale rotor 42 a has a wrap angle that is 6/7ths of the male rotor 42b. The wrap angle of the female rotor 42 a varies with the ratio ofnumber of lobes. The female rotor 42 a has a radius of about 2.5 inches(6.35 centimeters) and 7 lobes and the male rotor 42 b has a radius ofabout 3 inches (7.62 centimeters) and 6 lobes. The length of rotors 42is significantly smaller, on the order of about 20-30 percent smaller,than a conventionally sized screw compressor at the rated screwcompressor capacity. A person of skill in the art will appreciate thatanalytical techniques can be employed for other combinations of rotor 42profiles for a given rated screw compressor capacity within the scope ofthe present invention.

Employing a geometry/profile of rotors 42 for a screw compressor 12having a preselected screw compressor range and operable at anapproximately constant optimum peripheral velocity, allows for operationof the screw compressor 12 at 25 or more percent less than the ratedscrew compressor capacity without significant adverse rotor dynamiceffects. Screw compressor 12 has an improved rotor profile thatmaximizes internal flow area, internal friction due to relative motionof the rotor 42 surfaces is minimized, and leakage paths are reduced.This reduced leakage and higher flow tend to increase the screwcompressor 12 efficiency and reduce power wasted, which increasesoverall efficiency.

Referring now to FIG. 4, further details regarding an embodiment of thechiller 10 are presented. In particular, chiller 10 may include acontroller or controller system 60. Controller 60 may be arranged tocommunicate with the variable frequency drive 38, screw compressor 12,condenser 14 and evaporator 20. Chiller 10 may further include one ormore sensors. Sensors 66, 68, 70, 72 and 74, for example, may beemployed to sense and/or communicate torque, suction pressure and/ortemperature, discharge pressure and/or temperature, and/or othermeasurable parameter. Other sensors could be employed depending on theapplication in which screw compressor 12 is used. Signals 82 may becommunicated via wiring, fiber optics, wireless and/or a combination ofwiring, fiber optics and wireless. The sensors 66, 68, 70, 72 and 74communicate status signals 82 to controller 60 with data that areindicative of the operation of various components of the chiller 10.

The controller 60 may include processors, microcontrollers, analogcircuitry, digital circuitry, firmware, and/or software (not shown) thatcooperate to ultimately control operation of the screw compressor 12.The memory may comprise non-volatile memory devices such as flash memorydevices, read only memory (ROM) devices, electricallyerasable/programmable ROM devices, and/or battery backed random accessmemory (RAM) devices to store an array of performance relatedcharacteristics for the screw compressor 12. The memory may furtherinclude instructions which the controller 60 may execute in order tocontrol the operation of the screw compressor 12.

The controller 60 may receive status signals from one or more sensors66, 68, 70, 72 and 74 that provide information regarding operation ofthe screw compressor 12. Based upon the status signals, the controller60 may determine an operating mode and/or operating point of the screwcompressor 12 and may generate, based upon the determined operating modeand/or operating point, one or more command signals 62 to adjust theoperation of the screw compressor 12. The controller 60 may thengenerate command signals 62 that request the motor 52 to operateaccording to a preselected operating parameter(s) (e.g. a torqueprofile). For example, the controller 60 may enable operation at anoptimal torque and speed of screw compressor 12 to minimize losses,mechanical wear and losses. Further disclosure of a controller system 60suitable for use with embodiments of the present invention may be foundin co-pending application U.S. patent Ser. No. 12/544,582, assigned tothe assignee of the instant application, which is hereby incorporated byreference.

It should be apparent that variations on the control system 60 describedabove will be apparent to those skilled in the art. The control system60 may be implemented with electronic digital, analog, or a combinationof digital/analog control elements and low-voltage wiring. Otherconventional pneumatic tubing, transmitters, controllers, and relays arecontemplated.

In addition, it also will be readily apparent to one of ordinary skillin the art that the compressor system disclosed can be readilyimplemented in other contexts at varying scales. Use of various motortypes, drive mechanisms, and configurations with embodiments of thisinvention should be readily apparent to those of ordinary skill in theart.

Employing embodiments of the present invention, as compared toconventional approaches, increase full-load efficiency, yield higherpart-load efficiency and have a practically constant efficiency over agiven capacity range, controlled independently of power supply frequencyor voltage changes. Also, an advantage of embodiments of the presentinvention is that screw compressors 12 of different rated capacity caneach have a variable capacity and still have the approximately same thelevel of efficiency and without mechanical unloading.

Additional advantages include a reduction in the physical size of thescrew compressor and chiller system arrangement, improved scalability ofthe screw compressors throughout the operating range and a reduction intotal sound levels. Employing embodiments of screw compressor 12 canalso effectively reduce costs for the manufacturer, because it allowsfor one screw compressor at a rated screw compressor capacity (e.g.100-tons) to serve as an efficient screw compressor at a range ofpreselected screw compressor capacity range (e.g. 80 tons and 125 tons)without the need for multiple other screw compressors to be manufacturedat each additional target capacity within the preselected screwcompressor rated capacity range. Practically, embodiments of the presentinvention also allow for lower physical part count and inventory for aproduct family with no loss in capacity or performance due to powersupply because, for a given rated capacity of screw compressor (e.g.100-tons), the screw compressor 12 at 50 Hertz and 60 Hertz are nearlyidentical.

Embodiments of a bearing housing of a screw compressor are furtherdescribed. A screw compressor of a HVAC system may include one or morerotors. The one or more rotors of the screw compressor can be typicallysupported by bearings, such as for example axial and/or radial bearings.In some screw compressors, the bearings supporting the rotors can beenclosed and/or supported by a bearing housing.

FIG. 5 illustrates a partial side sectional view of a screw compressor100 that includes a first rotor 110 and a second rotor 120. The relativemotions of the first rotor 110 and the second rotor 120 can be used tocompress a working fluid, such as for example refrigerant vapor.

The first and second rotors 110 and 120 are positioned inside a rotorhousing 130. A bearing assembly including a bearing cover 170 and abearing housing 160 that is positioned at an axial end of the rotorhousing 130 along an axial direction defined by an axis C of the firstrotor 110. The bearing assembly generally covers the rotor housing 130at the axial end. The compressed working fluid is typically dischargedthrough the bearing assembly.

The first rotor 110 and the second rotor 120 include an axially extendedfirst shaft 112 and a second shaft 122 respectively along the axialdirection defined by the axis C of the first rotor 110. The first shaft112 and the second shaft 122 can be supported by one or more bearings,such as for example, discharge radial bearings 140 a and 140 b and/oraxial bearings 150 a and 150 b (sometimes referred to as “thrustbearings”) respectively. The discharge radial bearings 140 a and 140 band the axial bearings 150 a and 150 b can be attached to the first andsecond shafts 112 and 122. The axial bearing 150 a and 150 b, as well asthe discharge radial bearings 140 a, 140 b can be retained in the axialdirection by axial bearing retainers 142.

The bearings 140 a, 140 b, 150 a and 150 b can help support the firstrotor 110 and the second rotor 120 during operation, and facilitate therotation of the first and second rotors 110 and 120 around the axis C.For example, the discharge radial bearings 140 a and 140 b and the axialbearing 150 a and 150 b can help withstand forces that may be producedduring operation.

The bearing housing 160 is positioned at the axial end of the rotorhousing 130. The bearing housing 160 can be configured to enclose and/orsupport the discharge radial bearings 140 a and 140 b. The dischargeradial bearings 140 a and 140 b are typically positioned closer to therotor housing 130 than the axial bearings 150 a and 150 b in the axialdirection.

The bearing housing 160 includes a first cavity 162 and a second cavity164 configured to enclose and/or support the discharge radial bearings140 a and 140 b respectively. The bearing housing 160 can be a slab-likestructure, which may help provide a relatively uniform thermolexpansion. The discharge radial bearings 140 a and 140 b can besupported by walls of the first cavity 162 and the second cavity 164respectively, and may form interference fit with the walls of thecavities 162 and 164 respectively so that the bearing housing 160 cansupport the discharge radial bearings 140 a and 140 b during operation,such as for example, help the discharge radial bearings 140 a and 140 bwithstand forces that may be produced during operation.

The screw compressor 100 also includes the bearing cover 170, which canbe attached to the bearing housing 160 to form an enclosed space 180. Asillustrated in FIG. 5, the axial bearings 150 a and 150 b, portions ofthe first and second shafts 112 and 122, and the axial bearing retainers142 can be enclosed inside the space 180.

FIG. 6 illustrates an enlarged view of an area A of FIG. 5. Theslab-like bearing housing 160 includes the first cavity 162 and thesecond cavity 164 to enclose the first discharge radial bearing 140 aand the second discharge radial bearing 140 b respectively.

The first discharge radial bearing 140 a has a first length L1 and thesecond discharge radial bearing 140 b has a second length L2 in theaxial direction. The first cavity 162 has a first depth D1 and thesecond cavity 164 has a second depth D2 in the axial direction. Thefirst depth D1 can be configured to be about the same as the firstlength L1. In some embodiments, the first depth D1 can be configured tobe no more than the first length L1. The second depth D2 can beconfigured to be about the same as the second length L2. In someembodiments, the second depth D2 can be configured to be no more thanthe second length L2. The first depth D1 and the second depth D2generally are configured so that the first cavity 162 and the secondcavity 164 can enclose and/or support the first discharge radial bearing140 a and the second discharge radial bearing 140 b respectively. Insome embodiments, the first cavity 162 and the second cavity 164typically are configured not to enclose and/or support the first andsecond axial bearings 150 a and 150 b.

In some embodiments, the first discharge radial bearing 140 a can befully supported by the first cavity 162. In some embodiments, the seconddischarge radial bearing 140 b can be fully supported by the secondcavity 164. The term “fully supported” is referred to a situation wherethe support (e.g. interference fit) provided by the walls of the firstcavity 162 or the second cavity 164 to the first discharge radialbearing 140 a or the second discharge radial bearing 140 b respectivelywill stop increasing as the first depth D1 or the second depth D2increases.

The bearing housing 160 has a depth W1 in the axial direction. The depthW1 can have a minimal depth, which may be configured for example towithstand a compression pressure produced during operation.

In some embodiments, it can be beneficial to limit or minimize the depthW1, so that a mass of the bearing housing 160 can be limited orminimized. This can help limit or minimize, for example, thermalexpansion of the bearing housing 160, which may help maintain aclearance between the bearing housing 160 and the first rotor 110 or thesecond rotor 120.

The depth W1 can be affected by, for example, the minimal depth as wellas other factors such as the depth D1 and/or the depth D2. Generally,the larger the depth D1 and/or the depth D2 is, the larger the depth W1.In some embodiments, the depth D1 and/or the depth D2 may be configuredso that the first cavity 112 and/or the second cavity 122 can beconfigured to fully support the first radial bearing 140 a and/or thesecond radial bearing 140 b. In some embodiments, the depth D1 and/orthe depth D2 may be configured so that the bearing housing 160 does notextend to the axial bearings 140 a and/or 140 b in the axial direction.In some embodiments, the depth D1 and/or the depth D2 may be configuredto be no more than the length L1 and the length L2 respectively. In someembodiments, the depth D1 and/or the depth D2 may be configured so thatthe first cavity 112 and/or the second cavity 122 may be no more than adepth that is required to fully support the first radial bearing 140 aand/or the second radial bearing 140 b. In some embodiments, thedifference between the depth D1 and/or the depth D2 and the length L1and/or L2 may be about 3 mm, so that a maximum of about 3 mm of thefirst and/or second discharge radial bearings 140 a and 140 b is notenclosed and/or unsupported by the first and/or the second cavities 162and 164.

Referring to FIGS. 1 and 2, the bearing cover 170 generally has a domeshape, which can help define the space 180 with the bearing housing 160.The first and second axial bearings 150 a and 150 b can be enclosed inthe space 180. Referring to FIG. 6, the first axial bearing 150 a(and/or the second axial bearing 150 b, which is not illustrated in FIG.5) can have a clearance 182 between the first axial bearing 150 a(and/or the second axial bearing 150 b) and the bearing cover 170. Thefirst axial bearing 150 a and/or the second axial bearing 150 b does notgenerally need a support from the bearing housing 160 and/or the bearingcover 170 because the first axial bearing 150 a and/or the second axialbearing 150 b typically are configured to withstand forces in the axialdirection.

It is to be appreciated that because the bearing cover 170 can have theclearance 182 with the first axial bearing 150 a (and/or the secondaxial bearing 150 b), therefore the bearing cover 170 does not have tobe precision machined.

In the illustrated embodiment, the first axial bearing 150 a and/or thesecond axial bearing 150 b are arranged next to the first dischargeradial bearing 140 a and/or the second discharge radial bearing 140 b inthe axial direction. The first axial bearing 150 a and the second axialbearing 150 b are positioned further away from the first rotor 110and/or the second rotor 120 relative to the first and second dischargeradial bearings 140 a and 140 b respectively in the axial direction.Because the axial bearing 150 a and/or the second axial bearing 150 b donot necessarily require a support provided by the bearing housing 160 orthe bearings cover 170, it is not necessary to enclose and/or supportthe first axial bearing 150 a and/or the second axial bearing 150 b inthe first cavity 162 or the second cavity 164. To help minimize orlimit, for example, the mass of the bearing housing 160, the first depthD1 and/or the second depth D2 can be configured so that the first cavity162 and/or the second cavity 164 do not extend in the axial direction toenclose and/or support any portion of the first and/or second axialbearing 150 a, 150 b. The first axial bearing 150 a, the second axialbearing 150 b, and the axial bearing retainer 142 can be enclosed by thebearing housing 170.

It is noted that the arrangement of the first and second radial bearings140 a and 140 b, as well as the first and second axial bearings 150 aand 150 b are exemplary. The arrangement of bearings can be varied. Thenumber of discharge radial bearings and/or axial bearings can also vary.

FIG. 7 illustrates a known conventional bearing housing 360 of a screwcompressor 300. The bearing housing 360 includes a first cavity 362 anda second cavity 364. The first cavity 362 and the second cavity 364 areconfigured to enclose both a first and a second discharge radialbearings 340 a, 340 b, a first and a second axial bearings 350 a, 350 b,as well as axial bearing retainers 342. The bearing housing 360 iscovered by a bearing cover 370.

It can be seen for example that the bearing housing 360 is relativelylonger in an axial direction defined by an axis C3 of a first rotor 310relative to the bearing housing 160 of FIG. 6. That is, the bearinghousing 160 in FIG. 6 is relatively shorter in that the bearing housing160 encloses the first and second radial bearings 140 a, 140 b, but doesnot enclose the first and second axial bearings 150 a, 150 b or theaxial bearing retainers 142, whereas the bearing housing 360 in FIG. 7extends to enclose both the first and second discharge radial and axialbearings 140 a, 140 b, 150 a and 150 b, as well as the axial bearingretainers 342.

Screw compressor performance and reliability may be linked to theprecise location of the discharge radial bearings (e.g. the first andsecond radial bearings 140 a, 140 b), which can be configured to supportand locate the rotors (e.g. the first and second rotors 110 and 120). Byreducing the width W1 of the bearing housing 160 relative to theconventional bearing housing 360, the first and second cavities 162, 164can be produced more precisely and simply than the first and secondcavities 362, 364 of the convention bearing housing 360 due to shorterreaches and shorter machine boring bars. This can help position thefirst and second radial bearings 140 a and 140 b more accuratelycompared to the first and second radial bearings 150 a and 150 b, whichcan help both the performance and reliability of the screw compressor,by for example more accurately positioning the first and second rotors110, 120.

The bearing cover (e.g. the bearing housing 170) is typically not acritical part, and can be machined relatively cheaply and generally doesnot require precision machining tools. In the embodiment as disclosedherein, the bearing assembly includes the relatively long bearing cover170 and the relative short bearing housing 160 compared to therelatively short bearing cover 370 and the relatively long bearinghousing 360. The overall cost of making the bearing assembly is reducedcompared to a convention design. It is also simpler to make.

It is to be understood that it is possible to make the bearing housingthicker than it is necessary to accommodate the discharge radial bearingto it enclosed at lead a portion of the discharge radial bearings

With regard to the foregoing description, it is to be understood thatchanges may be made in detail, without departing from the scope of thepresent invention. It is intended that the specification and depictedembodiments are to be considered exemplary only, with a true scope andspirit of the invention being indicated by the broad meaning of theclaims.

What is claimed is:
 1. A bearing housing of a screw compressor,comprising: a cavity defined by the bearing housing, wherein a depth ofthe cavity is configured to be no larger than a length of a radialbearing of a screw compressor.
 2. The bearing housing of claim 1,wherein the bearing housing is configured to support a radial bearing ofthe screw compressor.
 3. The bearing housing of claim 1, wherein thebearing housing is configured to be positioned to cover an axial end ofthe screw compressor.
 4. A screw compressor, comprising: a rotorincluding a shaft; one or more radial bearings attached to the shaft;one or more axial bearings attached to the shaft; and a bearing housing,wherein the bearing housing defines a cavity that is configured tosupport the one or more radial bearings but not the one or more axialbearings.
 5. The screw compressor of claim 4, wherein the cavity has adepth and the one or more radial bearings have a length, and the depthof the cavity is not larger than the length of the one or more radialbearings.
 6. The screw compressor of claim 4, further comprising: abearing cover, the bearing cover configured to be attached to thebearing housing; wherein the bearing cover is configured to have aclearance with the one or more axial bearings.
 7. The screw compressorof claim 4, further comprising: a bearing cover, the bearing coverdefined a space; wherein the space defined by the bearing cover isconfigured to enclose the one or more axial bearings.
 8. The screwcompressor of claim 4, wherein the one or more radial bearings arepositioned closer to the rotor than the one or more axial bearings.
 9. Avariable capacity screw compressor comprising: a rotor housingcomprising a suction port, a working chamber, a discharge port, and atleast two screw rotors that include a female screw rotor and a malescrew rotor being positioned within the working chamber forcooperatively compressing a fluid; wherein the suction port, the atleast two screw rotors and the discharge port being configured inrelation to a selected rotational speed; the selected rotational speedoperates at least one screw rotor at an optimum peripheral velocity thatis independent of a peripheral velocity of the at least one screw rotorat a synchronous motor rotational speed for a rated screw compressorcapacity; a motor operable to drive the at least one screw rotor at arotational speed at a full-load capacity that is substantially greaterthan the synchronous motor rotational speed at the rated screwcompressor capacity; and a variable speed drive to receive a commandsignal from a controller and to generate a control signal that drivesthe motor at the rotational speed.